Parallel-flow type heat exchanger and air conditioner equipped with same

ABSTRACT

A parallel-flow type heat exchanger ( 1 ) is provided with two vertical-direction header pipes ( 2, 3 ), and a plurality of horizontal-direction flat tubes ( 4 ) that connect the header pipes with each other. The plurality of horizontal direction flat tubes are divided into a plurality of groups each comprising a plurality of flat tubes, and each of the groups constitutes a coolant path that lets coolant flow from one vertical-direction header pipe to the other. The upper limit for the number of flat tubes that constitute a coolant path for one turn is obtained from a prescribed numerical formula.

TECHNICAL FIELD

The present invention relates to a parallel-flow heat exchanger of a side-flow type, and to an air conditioner incorporating it.

BACKGROUND ART

A plurality of flat tubes are arranged between a plurality of header pipes such that a plurality of refrigerant passages inside the flat tubes communicate with the interior of the header pipes, with fins such as corrugated fins arranged between the flat tubes. So built are parallel-flow heat exchangers which are used widely in air conditioners for vehicles and in outdoor units of air conditioners for buildings, for instance.

An example of the structure of a parallel-flow heat exchanger is shown in FIG. 1. The top and bottom sides of FIG. 1 correspond respectively to the top and bottom sides of the heat exchanger. The parallel-flow heat exchanger 1 is of a side-flow type, and comprises two header pipes 2 and 3 extending in the vertical direction and a plurality of flat tubes 4 arranged between them and extending in the horizontal direction. The header pipes 2 and 3 are arranged parallel to each other, at an interval in the horizontal direction. The flat tubes 4 are arranged at a predetermined pitch in the vertical direction. At the stage of actual mounting in an appliance, the parallel-flow heat exchanger 1 can be installed at any angle to suit particular designs, and therefore, in the present description, the “vertical” and “horizontal” directions should not be interpreted strictly but be understood to merely give a rough notion of relevant directions.

The flat tubes 4 are elongate moldings of metal formed by extrusion and, as shown in FIG. 2, have formed inside them refrigerant passages 5 through which refrigerant is circulated. The flat tubes 4 are arranged with their longitudinal direction, i.e., the extrusion direction, aligned with the horizontal direction, and thus through the refrigerant passages 5, the refrigerant circulate in the horizontal direction. The refrigerant passages 4 comprise a plurality of refrigerant passages having the same cross-sectional shape and the same cross-sectional area arranged in the left-right direction in FIG. 2. Thus, in a vertical cross-sectional view, the flat tubes 4 look like a harmonica. The refrigerant passages 5 each communicate with the interior of the header pipes 2 and 3.

The flat tubes 4 have fins 6 fitted on their flat faces. Used as the fins 6 here are corrugated fins, but plate fins may instead be used. Of the fins 6 arranged in the up-down direction, the topmost and bottom most ones have side plates 7 arranged on their respective outer sides.

The header pipes 2 and 3, the flat tubes 4, the fins 6, and the side plates 7 are all made of metal with good thermal conductivity, such as aluminum. The flat tubes 4 are brazed or welded to the header pipes 2 and 3, so are the fins 6 to the flat tubes 4, and so are the side plates 7 to the fins 6.

The interior of the header pipe 2 is divided by two partitions P1 and P2 into three sections S1, S2, and S3. The partitions P1 and P2 separate the plurality of flat tubes 4 into three flat tube groups, so that a plurality of flat tubes 4 are connected to each of the sections S1, S2, and S3.

The interior of the header pipe 3 is divided by a single partition P3 into two sections S4 and S5. The partition P3 separates the plurality of flat tubes 4 into two flat tube groups, so that a plurality of flat tubes 4 are connected to each of the sections S4 and S5.

A refrigerant introduction/discharge pipe 8 is connected to the section S1, and another refrigerant introduction/discharge pipe 9 is connected to the section S2.

The parallel-flow heat exchanger 1 operates in the following manner. When the parallel-flow heat exchanger 1 is used as a condenser, refrigerant is fed into the section S1 through the refrigerant introduction/discharge pipe 8. The refrigerant that has entered the section S1 flows toward the section S4 through the plurality of flat tubes 4 connecting the section S1 to the section S4. This group of a plurality of flat tubes 4 constitutes a refrigerant path A. The refrigerant path A is indicated by a hollow arrow. Other refrigerant paths will be indicated by a hollow arrow each.

The refrigerant that has entered the section S4 turns back, then to flow toward the section S2 through the plurality of flat tubes 4 connecting the section S4 to the section S2. This group of a plurality of flat tubes 4 constitutes a refrigerant path B.

The refrigerant that has entered the section S2 turns back, then to flow toward the section S5 through the plurality of flat tubes 4 connecting the section S2 to the section S5. This group of a plurality of flat tubes 4 constitutes a refrigerant path C.

The refrigerant that has entered the section S5 turns back, then to flow toward the section S3 through the plurality of flat tubes 4 connecting the section S5 to the section S3. This group of a plurality of flat tubes 4 constitutes a refrigerant path D. The refrigerant that has entered the section S3 is then discharged out of it through the refrigerant introduction/discharge pipe 9.

In the present description, of the route traveled by the refrigerant, each segment between the refrigerant introduction/discharge pipe 8 or 9 and the next turning-back and between one and the next turning-back is referred to as “one turn.” Thus, the refrigerant paths A, B, C, and D each count as a one-turn refrigerant path.

When the parallel-flow heat exchanger 1 is used as an evaporator, refrigerant is fed into the section S3 through the refrigerant introduction/discharge pipe 9. Thereafter, the refrigerant travels in the reverse direction the route that it travels when the parallel-flow heat exchanger 1 is used as the condenser. Specifically, the refrigerant passes through the refrigerant path D, then the refrigerant path C, then the refrigerant path B, and then the refrigerant path A to enter the section S1, and is then discharged through the refrigerant introduction/discharge pipe 8.

In parallel-flow heat exchangers, elaborate designs have been made for improved performance. Examples are seen in Patent Documents 1 to 3 identified below.

In the parallel-flow heat exchanger described in Patent Document 1, inside a plurality of flat tubes that connect two header pipes together, a plurality of refrigerant passages with a fluidic diameter of 0.015 inches (about 0.38 millimeters) to 0.07 inches (about 1.78 millimeters) are formed parallel to one another. The outline of the cross section of those refrigerant passages is so designed as to have two or more comparatively straight portions that meet together and at least one dented portion formed where they meet. This design helps reduce the air-side front-face area obstructed by flat tubes, and thus makes it possible to increase the air-side heat transfer surface without increasing the air-side pressure drop.

In the parallel-flow heat exchanger described in Patent Document 2, refrigerant passages inside flat tubes are given a height of 0.35 millimeters to 0.8 millimeters. This helps reduce the sum of the drop in heat emission due to draft resistance and the drop in heat emission due to tube pressure loss, and thus helps improve heat emission performance.

In the parallel-flow heat exchanger described in Patent Document 3, a flow distribution parameter γ, i.e., the ratio of the resistance parameter β of flat tubes to the resistance parameter α of the refrigerant inlet-side header pipe is set at 0.5 or more. This helps prevent a concentrated flow of refrigerant through flat tubes connected to a refrigerant-inlet part of the header pipe where the pressure is higher. It is thus possible to make the pressure applied to the respective flat tubes even so that satisfactory flow distribution is achieved, and thereby to obtain satisfactory heat exchange performance.

LIST OF CITATIONS Patent Literature

Patent Document 1: JP-A-H5-87752

Patent Document 2: JP-A-2001-165532

Patent Document 3: JP-A-2000-111274

SUMMARY OF THE INVENTION Technical Problem

In a case where a parallel-flow heat exchanger is used as an evaporator, with respect to refrigerant passing through a refrigerant path, it is preferable that no such condition arise where more liquid refrigerant passes through some flat tubes and more gaseous refrigerant passes through other flat tubes; that is, it is preferable that no “uneven flow” occur. The present invention aims to provide a parallel-flow heat exchanger of a side-flow type that is designed optimally from the perspective of avoiding such an uneven flow with respect to the number of flat tubes constituting a refrigerant path. In particular, the present invention aims to optimize the number of flat tubes constituting a refrigerant path through which passes refrigerant with a large proportion of gaseous refrigerant.

Means for Solving the Problem

According to one aspect of the present invention, a parallel-flow heat exchanger of a side-flow type is provided with two header pipes extending in the vertical direction, and a plurality of flat tubes extending in the horizontal direction and coupling together the header pipes with each other. Here, the plurality of flat tubes are grouped such that each group comprises a plurality of flat tubes, each group constituting a one-turn refrigerant path through which refrigerant is passed from one to the other of the two header pipes extending in the vertical direction. Moreover, the upper limit of the number of flat tubes constituting the one-turn refrigerant path is determined to be within a range of ±2 of a value calculated using, when the parallel-flow heat exchanger is used in an outdoor unit of an air conditioner, the formula

n<3.0×10⁻⁴×Q+8.0,  (A)

and when the parallel-flow heat exchanger is used in an indoor unit of an air conditioner, the formula

n<4.2×10⁻⁴×Q+7.9,  (A)

where

n represents the number of flat tubes constituting the one-turn refrigerant path; and

Q represents rated capacity, given in watts (W). Used as Q is, for an outdoor unit, rated heating capacity and, for an indoor unit, rated cooling capacity.

When the parallel-flow heat exchanger configured as described above is used in an outdoor unit of an air conditioner, it is preferable that the lower limit of the number of flat tubes constituting the one-turn refrigerant path be determined using the formula

n>(αQ+β)×[(1.4×10⁻¹⁶)×L/(d×A′²)]^(0.5),  (B)

where

α=0.0161;

β=8.86;

d represents the hydraulic diameter, given in meters (m); and

A′ represents the refrigerant passage cross-sectional area of one flat tube, given in square meters (m²).

When the parallel-flow heat exchanger configured as described above is used in an indoor unit of an air conditioner, it is preferable that the lower limit of the number of flat tubes constituting the one-turn refrigerant path is determined using the formula

n>(αQ+β)×[(1.4×10⁻¹⁶)×L/(d×A′²)]^(0.5),  (B)

where

α=0.0228;

β=6.62;

d represents the hydraulic diameter, given in meters (m); and

A′ represents the refrigerant passage cross-sectional area of one flat tube, given in square meters (m²).

According to another aspect of the present invention, an air conditioner is provided with a parallel-flow heat exchanger configured as described above in an outdoor unit or in an indoor unit.

Advantageous Effects of the Invention

According to the present invention, it is possible to obtain a parallel-flow heat exchanger of a side-flow type that is free from an uneven flow depending on refrigerant circulation rate.

BRIEF DESCRIPTION OF DRAWINGS

FIG. 1 An outline configuration diagrams of a parallel-flow heat exchanger of a side-flow type;

FIG. 2 A sectional view along line II-II in FIG. 1;

FIG. 3 A table listing the specifications of flat tube samples;

FIG. 4 A table showing the correlation between refrigerant circulation rate and the uneven-flow-free number of flat tubes;

FIG. 5 A plot showing the correlation between refrigerant circulation rate and the number of flat tubes;

FIG. 6 A plot showing the correlation between cooling capacity and refrigerant circulation rate;

FIG. 7 A plot showing the correlation between heating capacity and refrigerant circulation rate;

FIG. 8 A plot showing the optimal range of the number of flat tubes for an outdoor unit of an air conditioner;

FIG. 9 A plot showing the optimal range of the number of flat tubes for an indoor unit of an air conditioner;

FIG. 10 A plot showing the correlation between refrigerant circulation rate and suction pressure;

FIG. 11 A plot showing the correlation between refrigerant circulation rate and the number of flat tubes;

FIG. 12 A plot showing the correlation between the number of flat tubes in an outdoor-unit heat exchanger and rated heating capacity;

FIG. 13 A plot showing the correlation between the number of flat tubes in an indoor-unit heat exchanger and rated cooling capacity;

FIG. 14 An outline configuration diagram of an air conditioner incorporating a parallel-flow heat exchanger according to the present invention, in heating operation; and

FIG. 15 An outline configuration diagram of an air conditioner incorporating a parallel-flow heat exchanger according to the present invention, in cooling operation.

DESCRIPTION OF EMBODIMENTS

A parallel-flow heat exchanger 1 of a side-flow type as shown in FIG. 1 wherein the number of flat tubes constituting a refrigerant path is set according to a method as described below is assumed to be a parallel-flow heat exchanger according to the present invention. The number of refrigerant paths, however, is not limited to four; more than four or less than four refrigerant paths may be provided.

First, the upper limit of the number of flat tubes 4 constituting a one-turn refrigerant path is determined; it is calculated, in a case where the parallel-flow heat exchanger is used in an outdoor unit of an air conditioner, using the formula

n<3.0×10⁻⁴×Q+8.0,  (A)

and in a case where the parallel-flow heat exchanger is used in an indoor unit of an air conditioner, using the formula

n<4.2×10⁻⁴×Q+7.9,  (A)

where n represents the number of flat tubes constituting a one-turn refrigerant path; and Q represents the rated capacity, given in watts (W).

Formula (A) was derived through experiments. The table in FIG. 3 lists the specifications of the flat tubes examined in the experiments. Sample a had a width of 16.2 mm, a thickness of 1.9 mm, and a refrigerant passage cross-sectional area of 13 mm². Sample b had a width of 13.9 mm, a thickness of 1.9 mm, and a refrigerant passage cross-sectional area of 11 mm² Sample c had a width of 16.2 mm, a thickness of 1.6 mm, and a refrigerant passage cross-sectional area of 11 mm² Sample d had a width of 19.2 mm, a thickness of 1.9 mm, and a refrigerant passage cross-sectional area of 14 mm².

The experiments were conducted in the following manner. Refrigerant was circulated through different numbers of flat tubes, and whether an uneven flow occurred was checked visually by thermography. For each of the four samples shown in FIG. 3, the refrigerant was circulated through it at varying circulation rates. The maximum numbers of flat tubes with which no uneven flow was observed at different circulation rates (in the present description, such a state is often referred to as uneven-flow-free) are listed in FIG. 4.

As will be seen from the table in FIG. 4, Sample a was used in Experiment 1. A refrigerant circulation rate of 27.3 kg/h gave a maximum uneven-flow-free number of 8. A refrigerant circulation rate of 42.5 kg/h gave a maximum uneven-flow-free number of 9. A refrigerant circulation rate of 64.3 kg/h gave a maximum uneven-flow-free number of 10. A refrigerant circulation rate of 63.2 kg/h gave a maximum uneven-flow-free number of 10.

Sample b was used in Experiment 2. A refrigerant circulation rate of 20.9 kg/h gave a maximum uneven-flow-free number of 9. A refrigerant circulation rate of 22.1 kg/h gave a maximum uneven-flow-free number of 8.

Sample c was used in Experiment 3. A refrigerant circulation rate of 59.2 kg/h gave a maximum uneven-flow-free number of 10. A refrigerant circulation rate of 48.8 kg/h gave a maximum uneven-flow-free number of 9. A refrigerant circulation rate of 26.4 kg/h gave a maximum uneven-flow-free number of 8.

Sample b was used in Experiment 4. A refrigerant circulation rate of 54.8 kg/h gave a maximum uneven-flow-free number of 8. A refrigerant circulation rate of 89.2 kg/h gave a maximum uneven-flow-free number of 8.

Sample d was used in Experiment 5. A refrigerant circulation rate of 26.6 kg/h gave a maximum uneven-flow-free number of 6. A refrigerant circulation rate of 44.3 kg/h gave a maximum uneven-flow-free number of 9. A refrigerant circulation rate of 67.3 kg/h gave a maximum uneven-flow-free number of 9.

FIG. 5 is a plot of the results of the experiments shown in FIG. 4. An approximation straight line is drawn, and from the approximation formula, the number of flat tubes is determined to be within a range of ±2 of the value given by

n=1.9×10⁻² m+7.8.  (a)

The refrigerant circulation rate m (kg/h) is typically set as a value proportional to the rated capacity of a product. How the refrigerant circulation rate correlates with the rated capacity is shown in FIGS. 6 and 7.

Using a rated heating capacity Q (in watts (W)), the refrigerant circulation rate m is given by

m=0.0161 Q+8.86.  (b)

Using a rated cooling capacity Q (in watts (W)), the refrigerant circulation rate m is given by

m=0.0228 Q+6.621.  (c)

The correlation between rated capacity and refrigerant circulation rate varies slightly from one product to another. Incidentally, the refrigerant circulation rate here is calculated in a simplified manner using the following formula:

(Refrigerant Circulation Rate m)=(Compressor Rotation Rate)×(Suction Pressure Density)×(Compressor Volume).

A parallel-flow heat exchanger, when used as an outdoor-unit heat exchanger of an air conditioner, functions as an evaporator in heating operation and, when used as a an indoor-unit heat exchanger of an air conditioner, functions as an evaporator in cooling operation.

Accordingly, as shown in FIG. 8, in a case where a parallel-flow heat exchanger is used as an outdoor-unit heat exchanger, using formulae (a) and (b) above, the upper limit of the number of flat tubes constituting a one-turn refrigerant path is determined to be

n=3.0×10⁻⁴ Q+8.0.

As shown in FIG. 9, in a case where a parallel-flow heat exchanger is used as an indoor-unit heat exchanger, using formulae (a) and (c) above, the upper limit of the number of flat tubes constituting a one-turn refrigerant path is determined to be within a range of ±2 of the value given by

n=4.2×10⁻⁴ Q+7.9.

This makes it possible to suppress an uneven flow.

Next, the lower limit of the number of flat tubes constituting each refrigerant path is determined As shown in FIG. 10, as the temperature at the outlet of the heat exchanger falls into the range

T_(out)<0° C.,

the suction pressure drops greatly; that is, the suction pressure drops sharply with respect to the refrigerant circulation rate. This is due to frost formation resulting from the outlet temperature falling below 0° C.

Let the temperature drop due to a pressure loss ΔP be T_(Dp), then

T_(Rin)−T_(Dp)<0° C.,

where T_(Rin) represents the inlet evaporation temperature of the refrigerant. The pressure loss ΔP is given in pascals (Pa).

That is,

P_(Rin)−ΔP>P_(lim),

where P_(Rin) represents the inlet evaporation temperature, and P_(lim) represents the saturation pressure of the refrigerant at 0° C.

Here,

ΔP=λ×L/d×ρ×u ²/2,

where λ represents the coefficient of friction between the inner wall of the flat tubes 4 and the refrigerant; L represents a tube path length, given in meters (m); d represents the hydraulic diameter, given in meters (m); ρ represents the refrigerant density, given in kilograms per cubic meter (kg/m³); and u represents the flow speed of the refrigerant, given in meters per second (m/s).

The flow speed is given by

u=M/ρA,

where M represents the refrigerant circulation rate, given in kilograms per second (kg/s); and A represents the sum of the refrigerant passage cross-sectional areas of the plurality of flat tubes constituting a one-turn refrigerant path, given in square meters (m²).

Thus,

ΔP=λ/2 ρ×L/dA ² ×M ².

Here, let the refrigerant passage cross-sectional area of one flat tube 4 be A′, then

A=nA′,

where n represents the number of flat tubes 4 constituting a one-turn refrigerant path.

Here,

ΔP<P_(Rin)−P_(lim).

Hence

λ/2 ρ×L/(dn²×A′²)×M²<P_(Rin)−P_(lim)

Here,

n²>M²×λ/2 ρ×L/dA′²×1/ (P_(Rin)−P_(lim)).

The above formula gives

n>M [λ/2 ρ×L/dA′²×1/(P_(Rin)−P_(lim))]^(0.5).  (d)

The refrigerant circulation rate m (kg/h), which is M as given in a different unit, is typically set as a value proportional to the rated capacity of a product; hence it can be expressed as

m=αQ+β.

How the refrigerant circulation rate correlates with capacity is shown in FIGS. 6 and 7. Using a rated heating capacity Q (in watts (W)), the refrigerant circulation rate m is given by

m=0.0161 Q+8.86.

That is, α=0.0161, and β=8.86.

Using a rated cooling capacity Q (in watts (W)), the refrigerant circulation rate m is given by

m=0.0228 Q+6.62.

That is, α=0.0228, and β=6.62.

For an outdoor-unit heat exchanger, rated heating capacity can be used; for an indoor-unit heat exchanger, rated cooling capacity can be used.

The correlation between rated capacity and refrigerant circulation rate varies slightly from one product to another. Incidentally, the refrigerant circulation rate here is calculated in a simplified manner using the following formula:

(Refrigerant Circulation Rate m)=(Compressor Rotation Rate)×(Suction Pressure Density)×(Compressor Volume).

On the other hand, it is common to keep the pressure loss below 200 kPa. Thus,

P_(Rin)−P_(lim)<200×10³.

The coefficient of friction λ varies with refrigerant circulation rate, refrigerant pressure, the shape of flat tubes, etc.; it is typically in the range of about 0.5 to about 0.05 in air conditioners for household use. The density p varies with refrigerant pressure and dryness; it is typically in the range of 20 to 70 kg/m³ with a gaseous refrigerant.

Thus, performing unit conversion from M to m gives

n>(αQ+β)×[π+L/(d×A′²)]^(0.5).

Here, π is given by

1.4×10⁻¹⁶<π<4.8×10⁻¹⁵.

In a case where the upper limit of the number of flat tubes calculated using formula (A) is exceeded by the lower limit of the number of flat tubes, it is preferable that the flat tubes be branched at the inlet or in the middle of the heat exchanger.

Here, considering that the pressure loss should be as low as possible, it is preferable to set II at its lowest value, namely 1.4×10¹⁶. Hence

n≧(αQ+β)×[(1.4×10⁻¹⁶)×L/(d×A′²)]^(0.5).  (B)

Thus, using formula (B), it is possible to determine the lower limit of the number of flat tubes constituting a one-turn path.

FIGS. 12 and 13 are plots of examples of the results of calculation using formula (B). FIG. 12 shows how the number of flat tubes in an outdoor-unit heat exchanger correlates with rated heating capacity. FIG. 13 shows how the number of flat tubes in an indoor-unit heat exchanger correlates with rated cooling capacity. These plots show the lower-limit values of the number of flat tubes constituting a one-turn refrigerant path as optimized according to rated capacity.

The parallel-flow heat exchanger 1 can be incorporated in a separate-type air conditioner. A separate-type air conditioner is composed of an outdoor unit and an indoor unit. The outdoor unit includes a compressor, a four-way value, an expansion value, an outdoor heat exchanger, an outdoor blower, etc. The indoor unit includes an indoor heat exchanger, an indoor blower, etc. The outdoor heat exchanger functions as an evaporator in heating operation, and functions as a condenser in cooling operation. The indoor heat exchanger functions as a condenser in heating operation, and functions as an evaporator in cooling operation.

FIG. 14 shows a basic configuration of a separate-type air conditioner that employs a heat pump cycle as a refrigerating cycle. The heat pump cycle 101 is composed of a compressor 102, a four-way value 103, an outdoor heat exchanger 104, a decompression-expansion device 105, and an indoor heat exchanger 106 connected in a loop. The compressor 102, the four-way value 103, the heat exchanger 104, and the decompression-expansion device 105 are housed in the cabinet of an outdoor unit. The heat exchanger 106 is housed in the cabinet of an indoor unit. The heat exchanger 104 is combined with an outdoor blower 107. The heat exchanger 106 is combined with an indoor blower 108. The blower 107 includes a propeller fan. The blower 108 includes a cross-flow fan.

The parallel-flow heat exchanger 1 according to the present invention can be used as a component of the heat exchanger 106 in the indoor unit. The heat exchanger 106 comprises three heat exchangers 106A, 106B, and 106C combined together like a roof covering the blower 108. The parallel-flow heat exchanger 1 can be used as any of the heat exchangers 106A, 106B, and 106C.

The parallel-flow heat exchanger 1 according to the present invention can also be used as the heat exchanger 104 in the outdoor unit.

FIG. 14 shows how heating operation proceeds. In this operation, high-temperature, high-pressure refrigerant is discharged from the compressor 102, and enters the indoor heat exchanger 106, where the refrigerant emits heat and condenses. The refrigerant then exits from the indoor heat exchanger 106, passes through the decompression-expansion device 105, and enters the outdoor heat exchanger 104, where the refrigerant expands as it absorbs heat from the outdoor air, before returning to the compressor 102. A current of air produced by the indoor blower 108 promotes heat emission by the indoor heat exchanger 106, and a current of air produced by the outdoor blower 107 promotes heat absorption by the outdoor heat exchanger 104

FIG. 15 shows how cooling operation or frost removal operation proceeds. In this operation, the four-way value 103 is so switched that the refrigerant circulates in the opposite direction compared with in heating operation. Specifically, high-temperature, high-pressure refrigerant is discharged from the compressor 102, and enters the outdoor heat exchanger 104, where the refrigerant emits heat and condenses. The refrigerant then exits from the outdoor heat exchanger 104, passes through the decompression-expansion device 105, and enters the indoor heat exchanger 106, where the refrigerant expands as it absorbs heat from the indoor air, before returning to the compressor 102. A current of air produced by the outdoor blower 107 promotes heat emission by the outdoor heat exchanger 104, and a current of air produced by the indoor blower 108 promotes heat absorption by the indoor heat exchanger 106.

It should be understood that the embodiment by way of which the present invention is described above is in no way meant to limit the present invention, which can thus be implemented with any modifications or variations made within the spirit of the present invention.

INDUSTRIAL APPLICABILITY

The present invention finds wide application in parallel-flow heat exchangers of a side-flow type.

LIST OF REFERENCE SIGNS

1 heat exchanger

2, 3 header pipe

4 flat tube

5 refrigerant passage

6 fin

7 side plate

A, B, C, D refrigerant path 

1. A parallel-flow heat exchanger of a side-flow type, comprising: two header pipes extending in a vertical direction; and a plurality of flat tubes extending in a horizontal direction and coupling together the header pipes with each other, wherein the plurality of flat tubes are grouped such that each group comprises a plurality of flat tubes, each group constituting a one-turn refrigerant path through which refrigerant is passed from one to the other of the two header pipes extending in the vertical direction, an upper limit of a number of flat tubes constituting the one-turn refrigerant path is determined to be within a range of ±2 of a value calculated using, when the parallel-flow heat exchanger is used in an outdoor unit of an air conditioner, the formula n<3.0×10⁻⁴×Q+8.0,  (A) and when the parallel-flow heat exchanger is used in an indoor unit of an air conditioner, the formula n<4.2×10⁻⁴×Q+7.9,  (A) where n represents the number of flat tubes constituting the one-turn refrigerant path; and Q represents rated capacity, given in watts (W).
 2. The parallel-flow heat exchanger according to claim 1, wherein the heat exchanger is used in an outdoor unit of an air conditioner, and a lower limit of the number of flat tubes constituting the one-turn refrigerant path is determined using the formula n≧(αQ+β)×[(1.4×10⁻¹⁶)×L/(d×A′²)]^(0.5).  (B) where α=0.0161; β=8.86; d represents a hydraulic diameter, given in meters (m); and A′ represents a refrigerant passage cross-sectional area of one flat tube, given in square meters (m²).
 3. The parallel-flow heat exchanger according to claim 1, wherein the heat exchanger is used in an indoor unit of an air conditioner, and a lower limit of the number of flat tubes constituting the one-turn refrigerant path is determined using the formula n≧(αQ+β)×[(1.4×10⁻¹⁶)×L/(d×A′²)]^(0.5).  (B) where α=0.0228; β=6.62; d represents a hydraulic diameter, given in meters (m); and A′ represents a refrigerant passage cross-sectional area of one flat tube, given in square meters (m²).
 4. An air conditioner comprising a parallel-flow heat exchanger according to claim 2 as an outdoor unit of the air conditioner.
 5. An air conditioner comprising a parallel-flow heat exchanger according to claim 3 as an indoor unit of the air conditioner. 